Mechanical torque converter



July 30, 1968 M. PRESTON 3,394,619

MECHANICAL TORQUE CONVERTER Filed May 19, 1967 7 Sheets-Sheet 1 IN VEN70R mar-fin. Prcszorz July 30, 1968 M, PRESTQN 3,394,619

MECHANICAL TORQUE CONVERTER Filed May 19, 1967 7 Sheets-Sheet 2 July 30,1968 M, ON 3,394,619

MECHANI CAL TORQUE CONVERTER Filed May 19, 1967 7 Sheets-Sheet 5 July30, 1968 M. PRESTON ECHANICAL TORQUE CONVERTER 7 Sheets-Sheet 4 FiledMay 19, 1967 Y L0 =ANGU1AP VEZOC/TY 0F 702G115 FMME FIG, 6

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July 30, 1968 Filed May 19, 1967 M. PRESTON MECHANICAL TORQUE CONVERTER7 Sheets-Sheet 5 July 30, 1968 M. PRESTON 3,394,619

MECHANICAL TORQUE CONVERTER Filed May 19, 1967 7 Sheets-Sheet 6 FIG. 3

July 30, 1968 M. PRESTON 3,394,619

MECHANICAL TORQUE CONVERTER Filed May 19, 1967 7 Sheets-Sheet 7 UnitedStates Patent M 3,394,619 MECHANICAL TORQUE CONVERTER Martin Preston,300 N. State St., Apt. 5701, Chicago, Ill. 60616 Continuation-impart ofapplication Ser. No. 594,061, Nov. 14, 1966. This application May 19,1967, Ser. No. 639,804

6 Claims. (Cl. 74-751) ABSTRACT OF THE DISCLOSURE A stepless,variable-speed power transmitting device in which the ratio of the inputand output shaft speeds depends (a) on the external torque load appliedto the output shaft and (b) on the speed of the power driven inputshaft. The transmission of power from the input to the output shaft isby means of a spinning rotor the kinetic energy of which undergoescyclical changes involving both the rotational speed and the massinertia of the rotor. During one phase of the working cycle energy istransmitted by a gear train from the input shaft to the rotor axle andduring another phase energy is transmitted from the rotor by gyroscopicforces to the output shaft.

This application is a continuation-in-part of my copending applicationSer. No. 594,061, filed Nov. 14, 1966, now abandoned, which in turn is acontinuation-in-part of my application, Ser. No. 501,637, filed Oct. 22,1965, now abandoned.

This invention relates to a variable speed, stepless, mechanical powertransmitting device which automatically adjusts the speed ratio betweenthe input and output shafts of the device both as a function of thespeed of the power source that drives the input shaft of the device andalso as a function of the external torque load that is imposed on theoutput shaft of the device.

In contrast to conventional hydrodynamic torque converters, the deviceoperates through the transfer of the momentum of rotating solid bodiesand its mechanical efiiciency depends solely on frictional losses ingears and bearings which can be minimized by the use of high quality,well lubricated, gears and antifriction bearings, while the efficiencyof the hydrodynamic torque converter is strongly affected by sizeableenergy losses due to turbulence and friction of the fluid used thereinfor the transfer of energy.

Another feature that contrasts this invention with the hydrodynamic typeare the performance characteristics which for the present device can besummarized as follows:

(a) If the output shaft is held fast, the torque acting thereon will beproportional to the square of the speed of the input shaft. Under theseconditions the power input would be zero were it not for the frictionallosses in the bearings and in the gearing.

(b) If the speed of the input shaft is held constant while the outputshaft performs work, the output torque either will vary linearly withthe change in the speed of the output shaft, or will remain constant.

Disregarding the effect of friction, these performance characteristicscan be defined concisely by the following mathematical relationships:

in which M is the input torque, M is the output torque, w is the speedof the input shaft and m is the speed of the output shaft. The constantsc and c depend on several built-in parameters. By the judicious choiceof 3,394,619 Patented July 30, 1968 these parameters the constant 0 cantake on positive or negative values, or can be made zero. In the lattercase the output torque will cease to be dependent on the output speedand thus the mechanism becomes a constant torque device suitable, e.g.,for driving with minimal loss of power a constant-tension windingmachine. The operating principle of the device is the utilization of theenforced precession of a spinning body whose axial moment of inertiachanges periodically in phase with the angle of precession and therebyproduces a unidirectional gyroscopic moment which serves as the outputtorque.

In the drawings:

FIGURE 1 is an elevational sectional view of a simple embodiment of thedevice.

FIGURES 1a and 1b are fragmentary sectional views of a modification ofthe embodiment shown in FIGURE 1.

FIGURE 2 is an elevational sectional view taken on line 22 on FIGURE 1.

FIGURES 2a, 2b and 20 show details of portions of FIGURE 2 taken onlines 2a-2a, 2b-2b and 2c2-c, respectively.

FIGURE 3 is a sectional plan view taken on line 3--3 on FIGURE 1.

FIGURE 4 is the view of an enlarged detail taken on line 44 on FIGURE 3.

FIGURE 5 is the view of an enlarged detail taken on line 5-5 on FIGURE3.

FIGURE 6 is a diagrammatic representation of part of FIGURE 2.

FIGURE 7 is a graph representing ceratin kinematic relations of theembodiment shown in FIGURES l, 2 and 3.

FIGURE 8 is an elevational view partly in section of a modifiedembodiment of the device.

FIGUR-E 8a is a fragmentary sectional view of a further modification ofthe embodiment shown on FIG- URE 8.

FIGURE 9 is an elevational sectional view taken on line 99 on FIGURE 8.

FIGURE 10 is a sectional plan view taken on line 1 010 on FIGURE 8.

FIGURE 11 is an enlarged detail of a component part of the embodimentshown in FIGURE 10 taken on line 111'1.

FIGURE 12 is an enlarged detail of a component part of the embodimentshown in FIGURE 10 taken on line 1212.

In the illustrations identical part numbers are assigned to componentsof similar design performing similar functions. In preparing theillustrations certain dimensional proportions were used which do notconform with an economical or practical design but Which appeareddesirable for better graphic representation.

A simple embodiment of the device shown in FIGURES 1-3, comprises aninput shaft 1 driven by an external power source and an output shaft 2delivering power to the outside, both shafts being journ-alled instationary housing 20 the top cover 20b of which carries gear case 17 inwhich the input shaft 1 is journalled, while the lower main body 20a ofthe housing is bushed for the output shaft 2.

The inside of the lower flanged portion of aforementioned gear case 17forms an internal gear which is in mesh with planet pinions 26 mountedon planet carrier 21 that is afiixed by pin,24 to input shaft 1 whoselower end is integral with bevel pinion 11. The sun gear 27 of thisplanetary gear set is provided with a tubular hub which is concentricwith shaft 1 and which carries on its lower end bevel pinion 7, keyedthereto.

When input shaft 1 is turned, the above described gearing arrangementcauses bevel pinion 7 to overrun bevel pinion 11, the former having morethan twice the rotational speed of the latter.

Immediately above, and concentric with bevel pinion 7 is stationaryanchor pinion 14 which is keyed to the lower end of the hollow hub-likeextension 20c of aforementioned housing cover 2017.

Torque-frame 12, consisting of top plate 12b, side plates 12c and baseplate 12a affixed to output shaft 2 by key 25, is rotatably supported inhousing 20 by bearing 31 at the top and by the bushing of shaft 2 at thebottom.

Said torque-frame 12 carries precession axle 8 afiixed thereto by pin 8b(FIGURE 1). Precession axle 8 in turn carries rotatably mounted bevelgears 4 (referred to as third bevel gear in the claims) and b (referredto as second bevel gear in the claims) and also ring gear 5a, the latterbeing supported by bearing 34 mounted in an annular recess of the web ofgear 5b. Each of gears 4, 5a and 5b are provided with two sets ofconcentric bevel gear teeth (FIGURES 1 and 3) the outer rows of whichbeing pair-wise in mesh with aforesaid pinions 14, 11 and 7,respectively.

Aforesaid precession axle 8, in addition to bevel gears 4 and 5b, alsocarries rotatably mounted spider hub 29 (FIGURES 2 and 3) having tworadial extentions 30 to which are bolted spider frames 6. Said spiderhub 29 and spider frames 6 serve as supports for two variableinertiarotors mounted symmetrically on opposite sides of said precession axle8.

Each one of said two rotors com-prises a spin-axle 3, an irregularhexagon shaped sleeve 23 which is affixed by pin 32 to said spin-axle 3,a multiplicity of fly weights 15 which are slidingly mounted on radialpins 16 and are forced outward by compression springs 18 (FIG- URES 4:and 5), and finally, thrust collar 13a which is slidingly mounted onspin-axle 3 and which has multiple fin-like extensions 13b serving ascams that engage cam rollers 19 (FIGURES 3 and 4) which are carried onpins 19a inserted across the central slot cut into fly weights 15(FIGURES 4 and 5).

Normally, the fly weights of the above described variable-inertia rotorare forced outward by said compression springs and by centrifugal forcesarising from the totation of the rotor-assembly. This condition isrepresented by FIGURE 4. However, if said thrust collar 13a is pressedtoward the fly weights, they will be forced inward (FIGURE 5) by theaction of cams 13b engaging cam rollers 19.

As stated before, said rotor-assembly is supported by spider hub 29 andspider frame 6, both of them having inserted bushings in which spin-axle3 is being journalled. The outer end of said spin-axle 3 carries bevelpinions 3a and 3b (FIGURES 2 and 3) through one-way clutches 33. Theseone-way clutches permit free unidirectional rotation of bevel pinions 3aand 3b relative to spin-axle 3 and in opposite direction to each other.Bevel pinion 3a meshes with the inner row of gear teeth of bothaforementioned bevel gear 4 and ring gear 5a, while bevel pinion 3bmeshes with the inner row of gear teeth of bevel gear 5b (FIGURE 3). Theposition of aforesaid thrust collar 13a (in the above descir-bed rotorassembly) which controls the position of said fly weights and therebythe variable moment of inertia of the rotor assembly, depends on theangular displacement of the spider assembly as it rotates aboutaforementioned precession axle 8, as can be seen in FIGURES 2, 2b and 3.Bell crank 10, pivoted on lug 22a which is integral with spider frame 6,carries roller 9 on its upper fork and rollers a on its bifurcated lower:fork. As roller 9 is forced inward by cam 8a which is cut into circularcam path 22 (FIGURES 1, 2, 2c and 3), rollers 10a will press againstaforesaid thrust collar 13a causing fly weights to move inward. Cam path22 is fixedly supported by welded lugs 28 which are bolted to sideplates 120 of aforementioned torque frame 12. The contour of the profileof cam 8a is such that the axial moment of inertia of each rotor isbeing kept at a constant minimum value during less than half arevolution of the spider frame and then, sequentially for an equalperiod, the moment of inertia of each rotor is kept at a steady maximumvalue. The transition between minimum and maximum values of the rotorinertia is being made gradually to obviate excessive inertia forces.

The angle 2A, shown in FIGURE 6 as subtending the ramp portions of theotherwise circular cam profile, represents the transitional phase of thevarying rotor inertia during a revolution of the spider assembly. The Yaxis shown thereon corresponds with the fixed axis of rotation oftorque-frame 12 and the Z axis represents the center line of precessionaxle 8. The angular velocity m is the precessional velocity of the spinaxis (which is the center line of spin axle 3) in the XY plane,furthermore, (.0 w and um, are the spin velocities of the rotor, bevelpinion 3a and bevel pinion 3b, respectively, all 'velocitiesbeingreferred to the XY plane. The total spin velocity of each of thesemembers actually includes the vector component of the angular velocity012 of torque-frame 12 (in the direction of the spin axis). The totalvelocities of the rotor and the two bevel pinions are denoted as 02 uand w respectively.

To describe the operation of the device it will be first assumed thatwhile input shaft 1 is driven at a constant angular velocity (.0 outputshaft 2 is held fast, that IS (U220.

Under these-.=conditions bevel gear 4 will remain at rest but ring gear5a and bevel gear 5b (FIGURE 1) will rotate together at the same speed.This will be so because the gear ratio between bevel pinion 11 (which isintegral with shaft 1) and its mating gear 5a is equal to the overallratio of the gear train interposed between shaft 1 and bevel gear 5b.Consequently, bevel pinions 3a and 3b (FIGURE 3), having the samecentral angle and being in mesh with ring gear 5a and bevel gear 51),respectively, will be driven in the same direction at identical speeds.It follows from the foregoing that spin axle 3 on which pinions 3a and3b are mounted (by means of one-way clutches 33) will be forced torotate about its own axis and also will be forced to precess, causingthe whole spider assembly to rotate about its own axis (FIGURES 2, 3 and6).

During the phase of increasing rotor inertia it will be only the one-wayclutch built into bevel pinion 3a which will be actively engaged intransmitting torque from this pinion to the spin axle, while bevelpinions 3b, whose one-way clutch :acts in the opposite direction, willnot participate in power transmittal.

Contrarywise, during the phase of decreasing rotor inertia, it will bepinion 3b that will transmit torque from the rotor to bevel gear 5b.Under these conditions the previously defined velocities of the rotorsand associated pinions will be constant and equal, that is,

Furthermore, w '=w "=w .Thus the kinetic energy absorbed by the rotorwhose inertia increases from a minimum to a maximum value will be fedback by the rotor whose energy decreases from a maximum to a minimumvalue. With no external work performed, the mean external torque appliedto input shaft 1 will be zero.

On the other hand, the gyroscopic moments acting on the spider frame(generated by the forced precession of the spin axes of the rotors) willnot balance during a full revolution of the spider frame about its ownaxis, owing to the cyclical change of the rotor inertia, as indicated inFIGURE 6. Considering the Y-component of the moments acting ontorque-frame 12, it can be shown that the instantaneous contribution ofthe individual rotor to this moment will equal: 'I w w sin where I isthe instantaneous axial moment of inertia of the rotor, (0 is its spinvelocity, w its precessional velocity and the angle between the Y axisand the spin axis. It can be, furthermore, deduced from FIGURE 6 thatsince during the first half of a revolution of the spider frame 1r) themean value of 1 will be greater than during the second half (1r 21r),the time integral of the Y'com-ponent of the gyroscopic moment will bepositive, that is, the mean moment applied to the torque frame will bein the direction of 00 as indicated in FIG- URE 6.

Designating the minimum value of the rotor inertia I; as I l and itsmaximum value as I -I-I' and, furthermore, since the effect of the meanvalve I of the inertia of one rotor cancels out that of the other, itwill sufiice to consider only the deviation of the instantaneous rotorinertia from its mean value as contributing to the moments acting on thetorque-frame.

Moreover, giving the profile of the ramp portion of cam 8a such acontour that the instantaneous value I of this deviation will be sin sinA s in the intervals -AA and (1rA)(1|-+A), while in the steady statephases 1:1, then the total instantaneous driving torque produced by thetwo rotors acting on the torque-frame can be expressed as the followingpositive quantity:

2l'w w lsin in which m and (a are expressible as linear functions of theinput speed w The graph of the variable I is shown in FIGURE 7.

It should be noted that the angle 2A subtending the ramp portions of thecam profile is not an arbitrary quantity but is definable as a functionof certain gear ratios in the device. It should be also pointed out thatthe net output torque available on output shaft 2 will include, inaddition to the above defined component of the gyroscopic moment, alsothe sum of the instantaneous moments contributed to the torque-frame bypinions 7, 11 and 14 drivingly engaged therewith.

In the foregoing the operating condition was described in which theinput shaft speed was constant and the output shaft was held fast. Next,the condition will be considered in which the input shaft speed, asbefore, is held constant but the output shaft speed varies. Thiscondition arises when the external load applied to the output shaftbecomes smaller than the stall torque (defined as the external resistingload just sufficient to stop rotation of the shaft). The output shaftthen starts rotating and its speed will vary as the external loadchanges.

Under these conditions and at increasing output shaft speed, bevel gears4 and 5b and ring gear 5a will all rotate in the same direction up to acertain output shaft speed, but bevel gear 5b will overrun ring gear 5a.Bevel gear 5b and ring gear 5a will rotate at the same speed (inclockwise direction if viewed from the right hand side in FIGURE 1) onlyif the output shaft is held fast, as heretofore described. However, assoon as the output shaft is released and starts rotating in thedirection indicated in FIGURE 1, the original (clockwise) rotationalspeed of both bevel gear 5b and ring gear 5a will be reduced and thespeed reduction of the latter will be greater than that of the former.Hence, comparing the resultant speed of these two gears it will beevident that bevel gear 5b, having sustained the smaller speed loss,will now run faster than ring gear 5a; that is, the former will overrunthe latter. How the rotation of the output shaft affects the rotationalspeeds of bevel gear 5b and ring gear 5a will become evident from FIGURE1 by the observation that torque frame 12 on which both bevel gear 5band ring gear 5a are journalled is rigidly connected to the output shaftand, therefore, the rotation of the output shaft in the directionindicated in FIGURE 1 will impose a counterclockwise rotation on bevelgear Sb because the latter is engaged with bevel pinion 7 and will alsoimpose a similar but more rapid counter-clockwise rotation on ring gear5:: as a result of the engagement of gear 5a with bevel pinion 11. Sincethe central angle is greater for bevel pinion 11 than for bevel pinion7, the counter-clockwise speed component is greater and the resultantnet clockwise speed is lesser for ring gear 5a than for bevel gear 5b.If the output shaft speed increases beyond that certain point, first,ring gear 5a will change direction of rotation, and then bevel gear 5b.The flow of power, will be similar to that described for the previouscase, that is, during the phase of increasing rotor inertia, power willbe transmitted from the input shaft through ring gear 5a and bevelpinion 3a to the rotor. During the phase of decreasing rotor inertia theflow of power will be from the rotor through bevel pinion 3b, bevel gear5b and through the planetary gear train to the input shaft. However, incontrast to the previous case, power will be transmitted from the inputshaft to the rotor during the period of constant rotor inertia. The netbalance of the time integrals of the flow of power to and from the inputshaft constitutes the work performed by the output shaft, fric tionbeing disregarded. The speed relations of the rotor and the therewithassociated bevel pinions 3a and 3b are shown in FIGURE 7.

The essential component of the device, the rotor comprising fiy weightsWhose rotary inertia is controlled by the angle of precession wasdescribed in the preceding text as consisting of weights slidinglyarranged on spindles. These weights could be also pivotally suspendedor, alternatively, arranged to roll on curved paths which arrangementmay have desirable features regarding dynamic balance. The presentedarrangement is merely an example whose graphic representation seemed tobe the most convenient compared to several alternative constructions.For the sake of compactness, the fly weights of the rotor shouldpreferably be made of heavy metal, such as tungsten alloy.

The constants c and 0 in the performance Equations 1 and 2 given earlierin the text, can be defined as follows:

tan a in which I, as previously defined, is the maximum deviation of theaxial rotor inertia from its median value,

a is half of the central angle of pinion 11,

B is half of the central angle of pinions 3a and 3b, 7 is half of thecentral angle of pinion 14,

6 is half of the central angle of pinion 7.

FIGURES la and 1b represent a slight modification of the originalembodiment shown in FIGURE 1. This modification involves the relocationof anchor pinion 14 from the top of the housing to the bottom andinvolves also a change in the connection between output shaft 2 andtorque-frame 12. In the original embodiment output shaft 2 is keyed tobase plate 12a of the torque-frame by key 25, whereas in themodification base plate 12d is journalled on a hub-like extension of thehousing and output shaft 2 is bolted to cam path 22 which in turn isrigidly connected to torque-frame 12.

Functionally, the effect of this modification is a reversal in thedirection of rotation of bevel gear 4. The effect of this change is alsothe sign reversal of the term tan 7 occurring in the above givendefinition of the constant 0 A more radical change of the originalembodiment is shown in FIGURES 8-12. Although the operating principle ofthis modified design is the same as that of the original embodimentshown in FIGURES 17, the performance characteristics of the modifiedembodiment offer a wider range of application. The difference betweenthe two designs can be summarized as follows:

(1) The latter design, having 8 instead of 2 rotors, delivers a moreuniform torque, obviating the need for flywheels.

(2) The latter design incorporates brake and clutch means that permitlimitation of the output shaft velocity, they provide also means forlocking together the input and output shafts and thereby cause theinternal mechanism of the device to rotate as a solid body with norelative movement between its component parts and finally, they make itpossible to lock up the drive completely. It should be noted, however,that aforesaid brake and clutch means could be incorporated in theoriginal embodiment shown in FIGURES l-7 with the same result.

As shown in FIGURES 8 and 9, housing 120 comprises cylindrical shell120a having a central hub extension at its bottom in which output shaft2 is journalled, cover 1201; whose upper cylindrical extension 120acarries bearing 140, brake disks 125 and cylindrical member 122 whichserves for setting aforesaid brake when downward force F is appliedthereto, said cylindrical member 122 being guided by pins 122a screwedinto said cylindrical extension 1200. The inner race of said bearing 140supports gear case 124 whose cover 134 carries mating brake disks 125 onsplines cut into the outer surface of its cylindrical extension and alsocarries clutch disks 126 on splines cut into the inner surface of itscylindrical extension. The lower flanged portion of said gear case cover134 forms an internal gear which is in mesh with planet pin-ion 137.Mating clutch disks 126 are carried on splines cut into planet carrier121 which is keyed to input shaft 1 by key 128. Said input shaft 1 iscoaxial with aforementioned output shaft 2 and carries on its lower endintegral bevel pinion 111. The lower end of input shaft 1 is journalledin pilot bearing 132 inserted into the extended top of output shaft 2.

Clutch pressure plate 131, slidingly mounted on the upper portion ofinput shaft 1, serves to engage clutch, and thereby lock up theplanetary gear set consisting of aforesaid planet carrier 121, planetpinions 127 and 137 and internal gear 117a. Said planet pinions 127 and137 are keyed to removable shaft 130, bushed in said planet carrier 121.The lower hollow cylindrical end of aforesaid internal gear 117a carriesbevel pinion 117, keyed thereto. Similarly, the lower hollow cylindricalextension of aforesaid gear case 124 carries anchor pinion 114, keyedthereto.

Torque-frame 112 consists of a yoke shaped lower part 112a, keyed tosaid output shaft 2 by key 129, of a top plate 112b, journalled bybearing 136 on a central hublike extension of housing cover 12012, oftwo lateral plates 112d (FIGURES 8, 9 and 10) and, finally, of two H-shaped, flanged, side plates 112e, adjacent to input shaft 1. Thesecomponents of the torque frame, being bolted together, form a rigidassembly which carries two precession axles 108, coaxially disposed andsymmetrically located with reference to input shaft 1. Each of theseprecession axles in turn carry a rotatable bevel gear 104a (near theirouter end), another rotatable bevel gear 105a (near their inner end) arotatable spider hub 139 (at their center) and two fixed cam disks 108a(at intermediate points, between said bevel gears and said spider hub).Attached to said spider hub 139 are four spider frames 106 bolted toradial, fin-like extensions of said spider hub. Bushings provided inaforesaid spider hub and spider frames serve as journals for four spinaxles 103, whose axes are set radially, spaced 90 degrees from eachother. Each of these spin axles carry on their outer end bevel pinions103b and 103a which are mounted thereon by means of one-way clutcheswhose free rotation is in opposite direction from each other.Additionally, spin axles 103 carry sliding fly weights 115 (FIG- URES l1and 12) mounted on pins 116 which are screwed into triangular sleeve 123aflixed by pin 133 to spin axle 103. Compression springs 118 tend toforce said fly weights outward. Cams 113b engaging ca-m roller 119 tendto force said fly weights inward. Said cams 11311 are integral parts ofbushed sleeve 113a, slidingly mounted on spin axle 103. Thrust collarwhich is provided with two cam followers 109 riding on aforementionedcam disks 108a, tends to force bushed sleeve 113a outward by means ofinterposed thrust bearing 135, causing fly weights to move radiallyinward. Said bevel pinion 103b meshes with ring gear 105b which iscarried on the outer circumference of aforesaid bevel gear 105a by meansof bearing 138. On the other hand, said bevel pinion 103a meshessimultaneously with aforesaid bevel gears 104a and 105a, the former ofwhich is drivingly connected with a concentric third bevel gear 104b.Bevel gears 105a and 104b and ring gear 105b in turn are pair wise inmesh with previously mentioned bevel pinion 111, anchor pinion 114 andbevel pinion 117, respectively.

Considering the operation of the device and assuming that brake is set(engaged) while clutch 126 is released (disengaged) it will be seen thatanchor pinion 114 will be locked to housing 120, while the planetarygear set (comprising planet carrier 121, planet pinions 127 and 137 andinternal gear 117a) will be fully operational. Under these conditionsthere will be functionally a consistent one-to-one relationship betweenthe corresponding parts of the device and that of the originalembodiment shown on FIGURES l7. Hence, both the functioning and theoperating characteristics will be as described previously.

Alternatively, assuming that brake 125 is released, while clutch 126 isengaged, it will be seen that since bevel pinions 111 and 117 and anchorpinion 114 will be locked together, the whole internal mechanismincluding the torque-frame and the therewith associated output shaftwill rotate together like a solid body. Hence, a straight drive-throughcondition will result.

If, as a third alternttive, both brake 125 and clutch 126 aresimultaneously engaged, the drive will be locked up altogether.

In the foregoing description item 132 was described as a pilot bearinginterposed between the ends of the input and output shafts. If thisbearing were replaced by a oneway clutch that would prevent the outputshaft from overrunning the input shaft, an automatic speed limitingfeature would be obtained. Similarly, if roller bearing 140 (interposedbetween the gear case and the housing) were replaced by a one-way brake,a certain limiting ratio of the output speed to input speed would beobtained.

The fragmentary detail shown in FIGURE 8a represents a furthermodification of the device affecting its operating characteristics, thatis, the value of the constants in the performance Equations 1 and 2.This modification brings bevel pinion 103b in mesh with bevel gear 104cwhile in the original embodiment it was bevel pinion 103a that wasmeshing with this bevel gear (designated as 104a in FIGURE 8).

In the illustrations all bevel gears are shown as being of thestraight-tooth-type. They could be replaced advantageously with spiralbevel gears.

It should be noted also that at the cost of narrowing the range of theperformance characteristics of the embodiment shown in FIGURES ll2, theanchor pinion (14 and 114) can be altogether eliminated from the device.In this case the mating outer row of gear teeth of bevel gear 4 (gear10412 in the latter embodiment) could be also omitted and the remainderof this gear (gear 104a in the latter embodiment) should be keyed to theprecession axle, this modified gear being referred to in the claims asfirst bevel gear. The aforesaid performance Equations 1 and 2 are validfor these modifications involving the elimination of the anchor pinions,but the value of the parameter 7, denoting one half of the central angleof the anchor pinions 14 and 114 should be taken as zero in theexpression defining the constant 0 occuring in the performanceequations.

What I claim is:

1. A mechanical torque converter comprising a stationary housing inwhich a first shaft and a'second shaft are journalled, the first shaftbeing driven by an outside source of power and the second shaftdelivering power to the outside; a rotatably mounted torque-framedrivingly connected to said second shaft, said torque frame beingprovided with a precession axle mounted on said torque-frame; a spiderframe rotatably mounted on said precession axle and carrying at leastone rotatably mounted spin axle the axis of which is at right angles tothe axis of said precession axle; a variable inertia rotor consisting ofat least two fly weights mounted movably on said spin axle in suchmanner that their radial distance from the axis of said spin axle can bevaried; means for controlling said radial distance as a prescribedfunction of the angle of rotation of said spider frame about saidprecession axle; a first and second pinion gear mounted on said spinaxle by means of one-way clutches whose free rotation is in oppositedirection from one another; gear means interposed between said firstshaft and said first and second pinion gears to transmit rotary motionfrom said first shaft to said spin axle causing it to rotate about itsown axis, and simultaneously to transmit rotary motion to said spiderframe causing it to rotate about said precession axle in such a mannerthat a unidirectional gyroscopic torque will be generated and applied tosaid torque frame.

2. A mechanical torque converter comprising a stationary housing inwhich a first shaft and a second shaft are journalled, the first shaftbeing driven by an outside source of power and the second shaftdelivering power to the outside; a rotatably mounted torque-framedrivingly connected to said second shaft, said torque frame beingprovided with a precession axle mounted on said torque-frame; a spiderframe rotatably mounted on said precession axle and carrying at leastone rotatably mounted spin axle the axis of which is at right angles tothe axis of said precession axle; a variable inertia rotor consisting ofat least two fly weights mounted movably on said spin axle in suchmanner that their radial distance from the axis of said spin axle can bevaried; means for controlling said radial distance as a prescribedfunction of the angle of rotation of said spider frame about saidprecession axle; a first and second pinion gear mounted on said spinaxle by means of one-way clutches whose free rotation is in oppositedirection from one another; gear means interposed between said firstshaft and said first and second pinion gears to transmit rotary motionfrom said first shaft to said spin axle causing it to rotate about itsown axis, and simultaneously to transmit rotary motion to said spiderframe causing it to rotate about said precession axle in such a mannerthat a unidirectional gyroscopic torque will be generated and applied tosaid torque frame, further characterized in that an anchor pinion isrigidly attached to said stationary housing and gear means areinterposed between said anchor pinion and said first pinion gear.

3. The device of claim 1 further characterized in that said first shaftinstead of receiving power from an Outside source, delivers power to theoutside and said second shaft instead of delivering power to the outsideis driven by an outside source of power.

4. The device of claim 1 wherein said gear means interposed between saidfirst shaft and said first and second pinion gears consist of a firstbevel pinion aflixed to said first shaft, a ring gear carried by saidprecession axle and provided with two concentric sets of gear teethwhich are in mesh with said first bevel pinion and with one of saidfirst and second pinion gears, respectively, a second bevel piniondriven through a gear train by said first shaft, and a second bevel gearcarried by said precession axle and provided with two concentric sets ofgear teeth which are in mesh with said second bevel pinion and with theother of said first and second pinion gears, respectively.

5. The device of claim 1 further characterized in that said first piniongear meshes with a first bevel gear keyed to said precession axle.

6. The device of claim 2, further characterized in that said gear meansinterposed between said anchor pinion and said first pinion gearconsists of a third bevel gear mounted on said precession axle andprovided with two concentric sets of gear teeth which are in mesh withsaid anchor pinion and said first pinion gear, respectively.

References Cited FOREIGN PATENTS 887,896 8/1943 France.

FRED C. MATTERN, JR., Primary Examiner.

T. C. PERRY, Assistant Examiner.

